Pinion bearing unit

ABSTRACT

A bearing unit, for supporting a pinion shaft having at one side a pinion head, includes a taper roller bearing and an angular ball bearing, the bearings having a common outer race ring and separate inner race rings. The inner race rings are spaced apart a distance along the pinion shaft, and the taper roller bearing has a contact angle α from 15° to 20° for absorbing mainly the radial force component of the tooth force at the pinion head, and the one-row angular ball bearing has a contact angle β of between 35° and 45° for absorbing mainly the axial force coming from the gear mesh and that induced by the taper roller bearing.

BACKGROUND OF THE INVENTION

The present invention refers to a hybrid pinion bearing unit for use ina motor vehicle front axle differential with negative hypoid offset andalso for motor vehicle power transfer units, and the purpose is toprovide an application-oriented optimization of the bearing assembly ofthe pinion shafts in drivelines. In relation to the present drivelines,due to the more and more increasing negative hypoid offset (i.e. therotational axis of the differential is situated below the correspondingpinion shaft) in the future, in front axle differentials and theconversion of the helical directions associated therewith of the taperdrive, the force conditions in the hypoid gearing will also change.Thereby the pinion axle in tractive operation in the future will bepulled into the crown wheel gears. In order to still be able to obtainsuch a low clearance of the flank of a tooth as possible, it is aboveall necessary to have a bearing assembly which is optimized forstiffness in axial direction.

Corresponding conditions are at hand at the so called PTUs (powertransfer units) for all-wheel driven vehicles, at which the drivenpinion axle via the crown wheel transfers its force to the drivendriving element, and consequently the axial force has the same effect asdescribed above.

In both cases the force conditions in the tooth engagement between crownwheel and pinion as compared to the conditions which are today common inrear wheel drives, differ thereby that the bearing unit according to thepresent invention is intended to make it possible to obtain an evidentimprovement regarding bearing stiffness with a simultaneous clearreduction of the effect of bearing losses.

U.S. Pat. No. 4,729,252 discloses a bearing unit for a pinion gear shaftincorporating a taper roller bearing and an angular ball bearing havinga common outer race ring and separate, axially spaced apart inner racerings, and it is stated that the contact lines of the taper rollerbearing intersect each other at an angle of less than 90° and that thecontact lines of the balls of the angular ball bearing likewiseintersect each other at an angle of less than 90°.

The stiffness of such a bearing combination is not optimal in view ofthe different characteristics for a taper roller bearing and an angularball bearing.

The purpose of the present invention is therefore to propose a modifiedbearing unit of this type, whereby the above problems are eliminated,and this is achieved by the invention disclosed and claimed hereinafter.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

Hereinafter, the invention will be further described by way of apreferred embodiment as illustrated in the accompanying drawing, inwhich:

FIG. 1 illustrates a cross-section through a portion of a pinion bearingunit according to the invention, and

FIG. 2 is a corresponding view of a pinion bearing unit as illustratedin FIG. 1, but connected to a pinion shaft.

DETAILED DESCRIPTION OF THE INVENTION

FIG. 1 shows in cross-section the upper half of a pinion bearing unitaccording to the invention, wherein a two-row bearing unit consists onthe pinion head side of a friction optimized taper roller bearing 1 andon the pinion tail side of an angular ball bearing 5. Both rows ofbearings have a common one-piece outer race ring 2, in which areprovided two opposite oil bores 3 (only one being shown), whereas theinner race rings 1 a and 5 a, resp. are axially spaced apart by aspacing sleeve 6. At the side of the common outer race ring 1 remotefrom the pinion head there is provided an external flange 4 for mountingpurpose. The bearings 1, 5 are mounted as preloaded bearings inback-to-back relation with an optimized preload with reference tostiffness and friction losses.

In order to obtain such a high bearing stiffness as possible and at thesame time a lower bearing friction and reduced effect loss also atextended bearing life span, it is necessary to design the internalbearing geometry in an ideal manner in relation to the requirements.

Thereby is it necessary to utilize the different characteristics of ataper roller bearing and of an angular ball bearing from applicationtechnical reasons for reaching the goal.

According to this, the taper roller bearing 1 in an ideal manner isoptimized for giving as low friction as possible for absorbing theradial force, which acts as a function of the torque to be transferredin the gears of the pinion head, in that taper roller bearing 1 has acontact angle α from 15° to 20°; for absorbing the axial force componentof the tooth force at the pinion head, and for absorbing the axial forceinduced by the taper roller bearing 1 at the same time the one-rowangular ball bearing 5 has a contact angle β from 35° to 45°.

A minimum of roller body reaction forces is experienced due to theaction obtained by the fact that the two contact angles are of differentsize, and this has a positive impact on bearing stiffness as well as onthe bearing friction losses. The large contact angle of the angular ballbearing provides for an increased carrying capacity and an optimizedaxial stiffness rate, respectively for the bearing unit.

For further reduction of loss effects and increase of the stiffness ofthe bearing it is also possible that the angular ball bearing isequipped with balls of ceramic material.

Dependent on the initially described force relationship at the pinionhead, the taper roller bearing 1 on the pinion head side is subjected toan axial unloading, which at high operational forces leads to a loadzone, limited to only a few taper rollers and a simultaneously essentialtilting between the inner and the outer race rings. In order to excludethe risk for impermissible high edge stresses at the rolling contact,the taper rollers are designed with a logarithmic profile at theenvelope diameter.

In order to fulfill the requirements for a bearing stiffness as high aspossible, at the same time as the bearing friction moment and thebearing effect losses are as low as possible, at the best possiblemanner in relation to the parameters of the bearing preload, the bearingunit is manufactured a) with consideration of usual fitting conditionsbetween the bearing outer race ring and the housing and between thebearing inner ring and the shaft respectively, and b) with an axialclearance before mounting of 0.03 to 0.07 mm. This leads to anapplication optimized bearing preload from 2000 to a maximum of 5000 N.

The increasing demand for low friction at simultaneous higher stiffnessdemand, which is the result of the demand for higher power density,thereby has been taken into account.

As the bearing unit comprising bearing outer race ring 2, inner racerings 1 a and 5 a and intermediate spacing sleeve 6 are not heldtogether before being assembled, it is shown in FIG. 1 how they fortransport purposes can be secured to each other by use of an innermounting sleeve 7 extending over a portion of each of the inner racerings 1 a and 5 a and over the intermediate spacing sleeve 6. Thedismounting force for the transport sleeve is limited to a maximum ofF=100 N.

Optionally it is possible to obtain a reduction of the constructionalheight tolerances, see 12 in FIG. 2, which leads to an evident reductionof the costs at the adjustment of the tooth flank clearance betweenpinion and crown wheel.

Further Possibilities:

For increasing the degree of integration it is proposed, to mount aradial shaft sealing ring 10 in the bore of the outer race ring on theside of the angular ball bearing. For use in a light metal housing isalso provided an O-ring 11 as a static oil seal. It is also conceivablewith the most different designs of a shoulder or a flange respectively,on the outer race ring for transfer of the axial forces to the housing.In FIG. 2 is illustrated the hybrid pinion shaft bearing unit in mountedcondition:

Mounting of the bearing unit on the pinion shaft 8 is limited to:

-   -   simultaneous pushing up or pressing up the two inner rings 1 a        and 5 a    -   rotating the pinion shaft 8 with the pinion head 8 a several        times until reaching engagement of the taper rollers against the        guide flange on the pinion head 8 a,    -   tightening the pinion nut 9 with an application specific        tightening torque in order to obtain the ideal pretension force.

The invention is not limited to the embodiment shown in the drawings anddescribed with reference thereto, but variants and modifications arepossible within the scope of the accompanying claims.

The invention claimed is:
 1. A bearing unit configured to support apinion shaft having a pinion head at one end, the bearing unit,comprising: a taper roller bearing and an angular ball bearing, thetaper roller bearing having a first inner race ring, the angular ballbearing having a second inner race ring spaced from the first inner racering by a spacer having a length, and the taper roller bearing and theangular ball bearing having a common outer race ring, wherein the taperroller bearing has a contact angle within a range of about 15° to about20° and the angular ball bearing has a contact angle within a range ofbetween about 35° and about 45°, and wherein a first end of the spaceris in direct contact with the first inner race ring and a second end ofthe spacer is in direct contact with the second inner race ring, andwherein the length is selected such the bearing unit has an axialclearance before mounting of 0.03 mm to 0.07 mm when the first innerrace ring, the spacer and the second inner race ring are in mutualcontact.
 2. The bearing unit as claimed in claim 1, wherein each of theangular ball bearing and the taper roller bearing has a plurality ofrolling bodies and the rolling bodies of at least one of the angularball bearing and the taper roller bearing are made of a ceramicmaterial.
 3. The bearing unit as claimed in claim 1, wherein rollers ofthe taper roller bearing have a logarithmic profile for reducing edgepressure.
 4. The bearing unit as claimed in claim 1, wherein the commonouter race ring has opposing axial ends and the bearing unit furthercomprises a sealing member fitted to the outer race ring on the endopposed to the pinion head.
 5. The bearing unit as claimed in claim 1,wherein the common outer race ring has an outer envelope surface and acircumferential groove in the outer envelope surface and the bearingunit further comprises a sealing member disposed within thecircumferential groove in the common outer race ring, the sealing memberbeing configured to provide a static oil seal against an inner wall of asurrounding housing.
 6. The bearing unit as claimed in claim 1,including an inner mounting sleeve securing the first and second innerrace rings and the spacer.
 7. The bearing unit as claimed in claim 1,including an inner mounting sleeve releasably securing the first andsecond inner race rings to the spacer.
 8. The bearing unit as claimed inclaim 1, wherein the axial clearance comprises an axial space between aroller of the taper roller bearing and the outer race ring or an axialspace between a ball of the angular ball bearing and the outer race ringor a combination of the axial space between the roller of the taperroller bearing and the outer race ring and the axial space between theball of the angular ball bearing and the outer race ring.